Variable displacement hydraulic assembly



July 2, 1963 1-. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLYOriginal Filed July 9, 1957 '7 Sheets-Sheet l S mm m 2 m m A M w I V m mX m H \X\ v T a "w E Q $3 w 5 a N a VW mm B mm b mN .TN m i I M/ ll H nll R 1 1 lh1l|al m S 3 H I. MN Q: I t .8 3 Ev mwm mm m w 2 mm M NN OH W.v? ni|| H 8 8 July 2, 1963 L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULICASSEMBLY 7 Sheets-Sheet 2 Original Filed July 9, 1957 S R. m O m R w A 0mm mmfiw A H .FLTW Q Q T E W mm om L 2 E Q Y \N. Mm B a m mm Fm 5 IN ,I1 IHMI d I m fl nnh m N m R: T l w l n u l I E. 2 3 mv mv NM d N WM mmNQ mm Om w a NM mm 2 m" E mm g .2 ON 3 \F I 0 o 0 0 o \L m m July 2,1963 1.. T. HARRIS 3,

VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Original Filed July 9, 1957 7Sheets-Sheet 5 ifiijmg g r INVENTOR.

LEE T HARRIS July 2, 1963 L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULICASSEMBLY 7 Sheets-Sheet 4 INVENTOR. LEE T HARRI 8 Original Filed July 9,1957 July 2, 1963 Y -r. HARRIS 3,095,708

VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Original Filed July 9, 1957 7Sheets-Sheet 5 INVENTOR.

LEE T. HARRIS July 2, 1963 L. T. HARRIS 3,095,703

VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Original Filed July 9, 1957XXII F19. 21 15 7 Sheets-Sheet 6 I I l INVENTOR.

a LEE T HARRIS July 2, 1963 Original Filed July 9, 1957 TO IGNITION 5W.

T0 BAND BRAKE SOLENOID SERVO-VALVE Fig.52

L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY 7 Sheets-Sheet 7TO ENGINE VACUUM TO BAND BRAKE SOLENOID SERVO-VALVE M LL w 120 BRUSHBRUSH ARM COMMUTATON ROTOR .TO TO RELAY R RELAY F INVENTOR.

BY LEET HARRIS United States Patent Q 3,095,708 VARIABLE DISPLACEMENTHYDRAULIC ASSEMBLY Lee T. Harris, 511 William St., Rome, N.Y. Originalapplication July 9, 1.957, Ser. No. 670,814, now Patent No. 3,044,409,dated July 17, 1962. Divided and this application Aug. 24, 1961, Ser.No. 141,555

9 Claims. (Cl. 60-53) This application is a division of my Patent3,044,409 dated July 17, 1962.

(This invention relates to hydraulic pump mechanisms and moreparticularly to a variable displacement pump mechanism that may be usedin a driving and driven configuration to comprise a unique hydraulictransmission system.

For some years now variable displacement pumps have been used inhydraulic transmission systems for coupling a prime mover to a load suchas in automobiles and other industrial applications. These transmissionsand pumps have generally been satisfactory for various applications buthave had certain disadvantages such as limited inputoutput speed ratiosrequiring special gear trains, high frictional and hydraulic lossesresulting in low efiiciency transmissions, costly and expensivemachining of highly stressed and critical parts resulting in high costlimited installations. According to the present invention I havediscovered a variable displacement hydraulic pump and transmissionsystem that permits infinite variation in input-output speed and torqueratios that is dynamically and statically balanced so as to minimizefrictional forces and that still may be manufactured largely by castingsin a cheap and simple manner.

Accordingly it is an object of the present invention to provide avariable displacement pump mechanism that is infinitely variable indisplacement. It is another object of the present invention to provide atransmission system that is infinitely variable in input-output speedratio throughout the design limits thereof. It is another object of thepresent invention to provide a variable pump mechanism that isdynamically and statically balanced so as to minimize frictional forcestherein. It is another object of the present invention to provide atransmission system that is extremely flexible in application andefficient in operation. It is another object of the present invention toprovide a transmission system wherein the driving pump may be locatedremotely from one or more driven pump mechanisms. It is another objectof the present invention to provide a variable displacement hydraulicpump mechanism that may be combined with a corresponding mechanism toprovide a transmission that rotates together at a one-to-one speed ratiowithout external gearing or locking. It is another object of the presentinvention to provide a transmission system that requires no gear trainor other speed control mechanism to connect it between a prime powersource and a load. It is a still further object of the present inventionto provide a variable displacement pump and transmission mechanism thatis highly efiicient, extremely simple to operate and economical toconstruct. These and other and further objects will be in part apparentand in part pointed out as the specification proceeds.

In the drawings:

FIGURE 1 is an axial section of the transmission with the rotatableinner portion shown at a position wherein the pitch axis of the drivingand driven pump divider disks are perpendicular to the plane of thedrawing and wherein Ice FIGURE 3 is a side elevation of the bearingmember in the same aspect relationship as shown in FIGURE 1;

FIGURE 4 is a top plan partially broken away of the valve and bearingmember of FIGURE 1;

FIGURE 5 is a vertical section of. the valve and hearing member alongthe line VV of FIGURE 3;

FIGURE 6 is an axial section of the ducting core which fits into thebearing member in the same aspect relationship as shown in FIGURE 4 withcertain parts shown in full lines for clarity;

FIGURE 7 is a view similar to FIGURE 6 taken on line VII-VII of FIGURE6.

FIGURES is a partial sectional view on line VIIIVIII of FIGURE 7;

FIGURE 9 is an end elevational view of the disk support assembly asviewed from the left hand side of FIGURE 2 and FIGURE 10;

FIGURE 10 is a side elevational view of the disk support assembly in thesame aspect relationship as shown in FIGURE 2;

FIGURE 11 is a top plan View of the disk support assembly of FIGURE 10;

FIGURES 12 and 12A are from left to right an end and side view of eitherof the two bushings that fit onto the transverse projections of the disksupport assembly;

FIGURE 13 is a diagrammatic drawing of the transmission main oil ductingdescribed by 360 counterclockwlse rotation as viewed from the left handend of FIG- URE 1 of the ducting core, starting at top center in respectto FIGURE 1;

FIGURE 14 is a side View of the pump chamber end piece in the sameaspect relationship as shown in FIGURE 1;

FIGURE 15 is a top plan view of the chamber end piece of FIGURE 14;

FIGURE 16 is a vertical transverse section of the pump chamber end piecealong the line XVIXVI of FIG- URE 14;

FIGURE 17 is a right half end view of the frontal end plate of the innerhousing as viewed from the right hand side of FIGURE 1;

the transmission is at an input-output forward speed ratio ofone-to-one;

FIGURE 2 is an axial section on line IIII of FIG-- FIGURE 18 is an axialsection of the frontal end plate of the inner housing in the same aspectrelationship as shown in FIGURE 1;

FIGURE 19 is an axial section of the frontal end plate of the innerhousing in the same aspect relationship as shown in FIGURE 2;

FIGURE 20 is an end View of the intermeshed rotor and divider disk ofthe driving pump as viewed from the right in FIGURES 1 and 2;

FIGURE 21 is a side view of the driving pump rotor and divider disk ofFIGURE 20 shown with the latter in section along line XXIXXI of FIGURE20;

FIGURE 22tis a partial transverse section of the drivmg pump ro or anddivider dis a XXII-XXII of FIGURE 21; k 310mg the hue FIGURE 23 is anenlarged view of a rotor vane and a partial circumferential section ofthe intermeshed divider disk showing the relationship of the two to eachother at a phase of the rotation cycle where the maximum angle ofincidence occurs in respect to the normal;

FIGURE 24 is a partial radial section of the dividerdisk along the lineO-XXIV of FIGURE 22 with the rotor removed;

FIGURE 25 isa partial radial section of the divider disk along the lineO-XXV of FIGURE 22 with the rotor removed;

FIGURE 26 is an enlarged exploded perspective view of adivider diskinsert and the corresponding urging spring;

FIGURE 27 is a schematic representation of a method for forciblymaintaining a relatively uniform speed relationship of the divider diskin respect to the rotor;

FIGURES 28, 29, and 30 are enlarged top, side, and end viewsrespectively of a rotor vane showing a further alternative method offorcibly maintaining a relatively uniform speed relationship of thedivider disk in respect to the rotor; and

FIGURES 31 and 32 are schematic drawings of a possible control systemfor automatically controlling the input-output speed ratio of thetransmission for automotive application.

Referring now to FIGURES l and 2, the transmission includes an inputshaft 1; an output shaft 2; an outer housing comprising end plate 3 andupper and lower casings 4 and 5 respectively; an inner housingcomprising end plate 6, front and rear barrel sections 8 and 9, centralsection 10, and rear section 12; ducting core 13; a first pump chamberbounded by pump chamber end piece 7, bearing member 11, and disk supportassembly 16-16; a second pump chamber bounded by member and rear section12; a driving pump including a rotor 14 and divider disk 15-15; a drivenpump or motor including rotor 17, and

divider disk 18-18; over-running clutch 19; and a band brake includingband 21 and drum 20.

The driving pump is infinitely variable in displacement from designmaximum forward output through zero output to design maximum reverseoutput as will be subsequently described. The rotor 14 is mechanicallycoupled to input shaft 1 and has therein six radial vanes 22 (see FIGUREcircumferentially spaced 60 apart at the centers with fiat sides andperipheral surfaces which are as surfaces of a common sphere whosecenter is the geometric center point of the pump chamber, and a coresection whose outer surface 23 is spherically contoured with the spherecenter coinciding with the same aforementioned geometric center point ofthe pump chamber, said surface 23 being the innermost surface orboundary of the pump chamber. Front face surface 24 of chamber end piece7, rear face surface 25 of bearing member 11 and peripheral surface 26of assembly 16-16 define the other principal boundaries of the firstpump chamber and are shaped so as to provide an oil sealing contact tothe end and peripheral surfaces of revolution of the vanes 22 of rotor14. Thus the six vanes 22 circumferentially divide the pump chamber intosix separate and equal volumes of fixed displacement.

Divider disk 15-15 comprises two main sections (see FIGURES 20, 21, and22); six slots 27 for accommodation of the vanes 22 of rotor 14,recesses 28 for accommodating inserts 29, valve chambers 30 foraccommodation of shuttle balls 31, oil channels 32, recess grooves 33,oil channels 34, and means for securely fastening the two main parts 15and 15 to each other. Said divider disk 15-15 intermeshes with rotor 14as shown in FIGURES 20, 21, and 22 and serves to longitudinally divideeach fixed volume between adjacent vanes 22 into two parts. Asvisualized in FIGURE 21 divider disk 15-15 would be assembled on rotor14 by sliding the part 15 onto said rotor 14 from the left and slidingthe other part 15 on from the right, one part, either 15 or 15containing the inserts 29 with corresponding springs and shuttle balls31, and the two parts 15-15 being joined by screws 36 as shown inFIGURES 20' and 24 or in some other suitable manner. Divider disk 15-15is supported in axial position by contact or proximity of the outer facebearing surfaces 37 of said divider disk 15-15 with the recessed matingsurfaces 37 of disk support assembly 16-16 (FIGURES 2 and 10) and inradial position by rotor 14. Disk support assembly 16-16 contains twoprojections, 33 and 38, which are journaled in mating bearing surfacescontained in the barrel sections 8 and 9 of the inner housing (FIGURE2), said disk support assembly 16-16 thereby being pivotally mounted ona transverse axis defined by an imaginary line connecting the twodiametrically opposite bearing centers, said line 4 intersecting thegeometric center of the pump chamber. Disk support assembly 16-16 issubject to control in its movements about the axis as will be describedherein.

In operation, oil sealing action between divider disk 15-15 and rotor 14is provided by contact or proximity of the spherically contoured innercontact surfaces 23 and outer slot contact surfaces 27 of said dividerdisk 15-15 with the mating spherically contoured contact surfaces 23 and22 of said rotor 14 in any supported position of said divider disk 15-15within the limits of movement of disk support assembly 16-16 about itstransverse axis. Oil sealing contact between the slots 27 and the radialsurfaces of vanes 22 is maintained by inserts 29 which are slidablymounted in recesses 28, said inserts 29 being continuously but yieldablyurged against vanes 22 by springs 35 or hydraulic pressure channeledinto recesses 28 by shuttle balls 31 or a combination of the two forcingmeans. Shuttle balls 31 actually serve two purposes: (1) by beingexposed to the pressure appearing on both sides of divider disk 15-15,between any two adjacent vanes 22, through orifices 39 said shuttleballs 31 are automatically seated against the orifices 39 which arefacing the lowest pressure, thereby leaving the orifices 39 which arefacing the highest pressure open, and permitting the highest availablepressure to be transmitted into valve chambers 30 and then into recesses28 via oil channels 32 wherein said high pressure acts against thebacksides of inserts 29 thereby counteracting the highest pressureappearing on the exposed face surfaces 40 of said inserts 29 (FIGURES 23and 26); (2) by leaving one orifice 39 open at all times, the oiltrapped in recesses 28 by inserts 29 is provided an escape when saidinserts 29 are forced back into said recesses 28 from extendedpositions.

In operation, it will be understood that when the input shaft 1 isrotated, as for example by a prime power source, that rotor 14 likewiserotates by virtue of its direct coupling to input shaft 1 and that sincedivider disk 15- 15 is mechanically intermeshed with rotor 14, it willlikewise rotate with said rotor 14 but will be permitted somecircumferential movement relative to rotor 14, said relative movementbeing restricted to the limits corresponding to the excess clearancebetween the face contact surfaces 40 of the opposing inserts 29, on eachside of a particular slot 27 when said inserts 29 are in their fullyretracted positions, as compared to the thickness of vanes 22. Saidexcess clearance should be at least as great as the maximum phasedeviation of corresponding points of rotor 14 and divider disk 15-15when calculated from the equation tan B=tan A cos P, in which P is thepitch angle of divider disk 15-15, A is any angle of the complete 360cycle of rotation of rotor 14 with either side of the pitch axis ofdivider disk 15-15 as the zero degree reference and B is the angle ofcorrespondence of the angle A on said divider disk 15-15 when translatedto the plane of rotation of rotor 14. Therefore, angle B subtracted fromthe angle A represents the phase difference of slots 27 in respect tocorresponding vanes 22 for various angles of rotation and for variouspitch angles at which divider disk 15-15 may be set between zero andmaximum design limits. In order to arrive at the minimum designclearance for vanes 22 in slots 27, angle A should be taken at angles of45, 225 and 315, and angle P at the maximum design pitch angle ofdivider disk 15-15'. Maximum angle P is 15 in the illustrated embodimentwhich causes the maximum phase difference between rotor 14 and dividerdisk 15-15 to be approximately :1"; therefore the total clearancebetween face contact surfaces of opposing inserts 29 when in their fullyretracted positions in grooves 28 should be at least equal to thethickness of vanes 22 plus r sin 2, where r is the radius from thegeometric center of the pump housing.

Although inserts 29 will perform most of the yielding to the phasevariation between rotor 14 and divider disk 15-15', the frictionalresistance to rotation against divider disk -15 will tend to cause saiddivider disk 15-15 to lag behindrotor 14 to the limits imposed by theexcess clearance between opposed inserts 29 which will tend to causecyclic variations in the speed of said divider disk 15-15 in respect tothe speed of rotation of rot-or 14. FIGURE 27 illustrates a meanswhereby a more nearly constant speed rotation of divider disk 15-15 inrespect to rotor 14 may be assured, when the transmission is operatingat an input-output speed ratio of one to one. A groove 41 is provided inthe peripheral surface 22 of each vane 22, the sides of which are shapedto provide contact to at least one point of a mating projection 42,located at the peripheral center of each slot 27 in divider disk 15-15,at all phase angles of rotation between 315 to 45 and 135 to 225, withthe zero degree reference considered to be in coincidence with one sideor the other of the transverse pitch axis of divider disk 15-15.

A further alternative means whereby more nearly constant speed rotationof divider disk 15-15 in respect to rotor 14 may be assured (FIGURES 28,29, and 30) is to shape the sides of vanes 22 so as to keep one or bothof the opposing inserts 29 in any one slot 27 fully retracted into theirrespective recesses 23 at all phases of cyclic rotation between theangles of 315 to 45 and 135 to 225 with the zero degree reference againchosen to coincide with either side of the transverse pitch axis ofdivider disk 15-15.

By inspection of the drawings, it will be apparent that divider disk15-15 operating in a complementary manner with rotor 14 essentiallydivides the pump chamber into two pumping sections (one of each side ofdivider disk 15-15) which operate in'a diametrically opposite fashion toeach other with each section having its own input-output ports,represented by the numbers 43 and 43 for the front pump section andnumbers 44 and 44 for the rear pump section as shown in FIGURE 4 and asshown schematically in FIGURE 13. The ports 43, 43, 44 and 44 shouldeach have a circumferential length of approximately 120, leaving anuncut-away circumferential section of approximately 60 in length betweeneach port which will serve to trap the fluid between ad-' jacent vanes22 at maximum of minimum points as the case may be and thereby toprevent unrestricted circulation of oil between input and output portswhich would tend to bypass the pump. If the rotor employs more or lessthan six vanes as shown in the drawings, the ports should of course beof appropriate circumferential length corresponding to the number ofvanes.

Referring now to FIGURE 1 it is obvious that when the plane or rotationof divider disk 15-15 is completely vertical (zero pitch angle)represented by alignment with the letters N-N, the volume of fluidentrapped between adjacent vanes 22 and the surfaces of the pump chamberremains constant when rotor 14 and divider disk 15-15 rotate and thattherefore all of the fluid contained within the chamber rotates with therotor with no fluid input or output through ports 43, 43, 44 and 44. Itshould be equally obvious that if the pitch angle of divider disk 15-15is at any valve other than zero that the individual fluid volumesbetween adjacent vanes 22 on each side of divider disk15-15 will vary ina sine manner as motor 14 progresses from 0 through 360 for eachrotation cycle, the total changebetween minima and maxima volumes beinga direct function of the pitch angle of divider disk 15-15. Therefore alow compressibility fluid such as oil when contained therein will beforcibly circ-ulated into and out of the pump chamber through ports 43,43, 44 and 44 in amounts proportional to the pitch angle of divider disk15-15. In the illustrated embodiment, maximum forward circulation isachieved when the plane of rotation of divider disk 15-15 is alignedwith the letters F-F (FIGURE 1); i.e. the pitch angle of divider disk15-15 isat the forward maximum of 150. Likewise, the maximum 6 reversecirculation will be achieved when the plane of rotation of divider disk15-15 is aligned with the letters R-R. Specific circulation paths forthe fluid will be described in detail herein.

Examination of the underlying principles of operation will disclose thatinsofar as rotor 14 is concerned, no appreciable unb-alanced forces inrespect to the housing occur but that the hydraulic forces on dividerdisk 15-15 would be severelyunbalanced if not compens-ated for. It isfor said purpose of compensation that recess grooves 33 are providedaround each peripheral side of divider disk 15-15 (FIGURES 20' and 21)with oil channels 34 providing hydraulic communication between saidrecess grooves 33 and the opposite side of divider disk 15-15 in thepump chamber. The effective area of recess groove 33 is chosen so thatsaid area multiplied by the radius of the effective center of fluidpressure of said recess groove 33 is equal to the effective area of thepart of divider disk 15-15 contained inside the pump chamber betweenadjacent vanes 22 multiplied by the radius of the efiiective center offluid pressure of said second area. In this way any hydraulic forcesappearing on the inside areas of divider disk 15-15 will be dynamicallycounteracted by proportionate forces created by transmission of pressurethrough oil channels 34 to recess grooves 33 located on the oppositesides. The obvious purpose of providing hydrostatic balance for dividerdisk 15-15 is to reduce the frictional resistance to rotation of saiddivider-disk 15-15 in disk support assembly 16-16 occurring as a resultof hydraulic loading, and thereby to enhance the efliciency of powertransfer and decrease mechanical wear. Another method for reducing thefriction, in lieu of hydrostatic balance as described above, is tosupport divider disk 15-15 on ball or tapered roller bearings withindisk support assembly 16-16. A third method might be a combination ofthe above two methods.

The driven pump or motor, hereinafter to be referred to as the drivenpump, in the illustrated embodiment is essentially identical to thedriving pump except that divider disk 18-18 has a fixed plane (pitchangle of 15) of rotation about its axis by virtue of its being supportedin the inner housing and accordingly is a fixed displacement pump ormotor. It will be apparent, however, that the driven pump may be of thevariable displacement type also and that said driven pump whether of thefixed or variable displacement type may :be similar to or different fromthe driving pump in basic design. The rotor 17 with six vanes 22A,divider disk 18-18 and associated pump chamber acting in cooperationwith each other form the essential elements of the driven pump, theoperation of which will be fully understood by reference to the abovedescription of the driving pump. 7

Ducting core 1-3 provides means of fluid communication between thedriving pump and driven pump by providing fluid channels '45 and 45 asshown in FIGURES 1 and 2 and as shown schematically in FIGURE 13.Channel 45 connects ports 43' and 44 of the driving pump to ports 436and 47 of the driven pump and channel 45 connects ports 43 and 44 of thedriving pump to ports 46 and 47 of the driven pump.

It will be apparent that when input shaft 1 is. rotated clockwise whenviewing FIGURES 1 and 2 from the left, and that when divider disk 15-15is set at a pitch angle greater than zero toward alignment with theletters F-F of FIGURE 1, fluid will be expelled from the chamber of thedriving pump through ports 43 and 44 into channel 45 from which saidfluid will be forced into the chamber of the driven pump through ports46* and 47 thereby causing rotor 17 to rotate in the same clockwisesense as rotor 14, causing said rotor 17 at the same time to expel alike quantity of fluid from the chamber of the driven pump through ports46' and 47 into channel 45 from which said like quantity of fluid willbe returned to the chamber of the driving pump through ports 43 and 44.Fluid flow under the above conditions is shown schematically by thesolid arrows in FIGURE 13. When divider disk 15-15 is set at a pitchangle greater than zero toward alignment with the letters R-R of FIGURE1 and rotation of input shaft '1 is in the same clockwise sense, theflow of fluid is reversed and would be as shown schematically by thedashed arrows in FIGURE 13, in which case output shaft 2 would rotatecounter-clockwise or in the opposite direction to the direction ofrotation of input shaft \1.

It is also apparent that regardless of the direction of rotation ofinput shaft 1, with divider disk 15-15 pitched toward alignment with theletters F-F in FIGURE 1, the output shaft 2 will rotate in the samedirection as input shaft 1, and with divider disk 15-15 pitched towardalignment with the letters R-R in FIGURE 1, rotation of output shaft 2will be opposite to that of input shaft 1.

It will be understood that the ratio of speed of the input shaft 1 tothe speed of the output shaft 2 will be inversely proportional to theratio of the displacement of the driving pump to the displacement of thedriven pump and that the output-input torque ratio will be directlyproportional to the input-output speed ratio, neglecting frictionallosses. In the illustrated embodiment of the transmission, a one to oneforward ratio of speed between input shaft 1 and output shaft 2 occurswhen the plane of rotation of divider disk 15-15 is in alignment withthe letters F-F in FIGURE 1.

Pitch control of divider disk 15-15 is effected by control of themovement of disk support assembly 16- 16' about its transverse axis bypermitting oil under pressure to enter either control chamber A throughchannel 48 or control chamber B through channel 49 and to exit from theother (see FIGURE 1). Disk support assembly 16-16 has projections orvanes 50 (FIGURES 9, 10, 11) which together with cylindrical surfaces 51(FIG- URES 17, 18, 19) in end plate 6 and the spherical peripheralsurfaces 51' of control chambers A and B (FIGURE l1) forms a fluid sealto prevent the unrestricted flow of oil between control chambers A andB. The axial center line of said cylindrical surfaces 1 coincides withthe transverse axis of rotation of disk support assembly 16-16. Vanes 52are similar to vanes 50 but are not necessarily designed to provide oilsealing contact with any surfaces of the housing, the essential purposeof said vanes 52 being to balance the weight of vanes 50 in respect tothe longitudinal axis of the transmission when the inner housing isrotating, although said vanes 52 also serve to strengthen the disksupport assembly 16-16 at points where needed.

To increase the input-output forward speed ratio or to decrease theinput-output reverse speed ratio, oil under pressure is permitted toenter control chamber B through channel 49 from the control pressuresource and a like amount of fluid is permitted to exit from controlchamher A through channel '48 and to return to the sump or inlet of thecontrol pressure source, which action causes disk support assembly 16-16to move in the direction toward alignment of the transverse center planeof said divider disk 16-16 with the letters F-F of FIGURE 1. To decreasethe input-output forward speed ratio or to increase the input-outputreverse speed ratio, oil under pressure is permitted to enter controlchamber A through channel 48 and to exit from control chamber B throughchannel 49 and to return to the sump or inlet of the control pressuresource, which action causes disk support assembly to move in thedirection toward alignment of the transverse center plane of saiddivider disk 16-16 with the letters R-R of FIGURE 1. Annular grooves 48and 49' are provided in end piece 3 of the outer housing as a means ofeffecting hydraulic communication between a control system locatedexternally to the inner housing and channels 48 and 49 respectivelyunder conditions when the inner housing is rotating as well as when saidinner housing is held motionless in re- 8 spect to the outer housing asshown in FIGURES 1 and 2.

In order to obtain a mechanical indication of the transmissioninput-output speed ratio which might be required for a follow-up type ofcontrol system, a ratio indicating linkage is provided as shown inFIGURE 6. This linkage starts with a pair of gears 53 and 54 fixed onshaft 55 and mounted so that gear 53 meshes with a circular gear segment56 (FIGURE 11) on disk support assembly 116-16 and gear 54 meshes withgear rack 57 mounted in sleeve 58 inside dusting core 13. When disksupport assembly 16-16 rotates about its transverse axis a proportionateangular rotation, corresponding to the gearing ratio between said gear'53 and said circular gear segment 56, is induced in gear 54 causinggear rack 57 to move longitudinally together with sleeve 58 to which itis mechanically fixed. Shaft 59 is threaded into sleeve 58 at one end(and therefore moves longitudinally with said sleeve 58) and has fixedin position on the other end of shaft 59 by retaining nut 61 a bearingassembly which is slidably mounted in output shaft 2 (FIGURE 2). Anindicating sleeve 62 having thereon a rim 66 is slidably mounted onoutput shaft 2 and mechanically fastened to bearing assembly 60 byscrews 63 which are free to move longitudinally in slots 64, saidindicating sleeve 62 thereby being caused to move longitudinally as aunit with bearing assembly 60. A lever assembly consisting of two arms65 and 65, the extremities of which bear against rim 66 of indicatingsleeve 62 fixed on a shaft 67 to which a third arm 68 is rigidlyfastened extends the linkage away from the output shaft to the controlmechanism. Shaft 67 is trunnioned on a transverse axis so that itscenter line is the pivot point for the lever assembly which is caused topivot about its axis to follow the movement of indicating sleeve 62 byspring 85 or other equivalent means. A rod 69 is pivotally mounted onthe end of arm 68 so as to translate the rotative movement of said arm68 into longitudinal motion for utilization by a transmission controlmeans. Rotative motion could just as well be imparted to thetransmission control means by appropriate gearing arrangement from shaft67 of the lever assembly if said rotative motion were more suitable foruse by the control means.

When disk support assembly 16-16 rotates about its transverse axis aproportionate angular rotation, corresponding to the gearing ratiobetween said gear 53 and said circular gear segment 56, is induced ingear 54 musing gear rack 57 to move longitudinally together with sleeve58 to which it is mechanically fixed. This movement is in turntransmitted through the linkages described until rod 69 moves theindicating mechanism to the corresponding indication.

Bypass control piston 70 (FIGURES 6 and 7) forming the end of shaft 59which attaches to sleeve 58, contains an annular groove 71, the purposeof which is to bypass a nominal amount of oil from the high pressurechannel in ducting core 13 via orifices 72 and 72 to the low pressurechannel in said ducting core 13 when a load connected to output shaft 2is being accelerated from the motionless state in either the forward orreverse direction. Bypass control piston 70 acts in cooperation withorifices 72 and 72' to form a progressively smaller restriction to thebypass oil path as divider disk 15-15 is increased in pitch in eitherthe forward or reverse direction from the zero pitch angle up to apredetermined pitch angle at which point the bypass path will becompletely obstructed by the smooth cylindrical surface of the bypasscontrol piston. By this means, torque in varying degrees may be appliedto a motionless load without stalling the power source, therebypermitting smoother starts. For some forms of application of thetransmission a bypass valve obviously would not be required.

Shuttle valve 73 (FIGURES 6 and 7) may be mounted in the ducting core 13by threaded means as shown and comprises a valve housing 74 and shuttlespool 75. The

valve on one end is exposed to fluid channel 45 which is normally thehigh pressure channel and on the other end is exposed to fluid channel45' which is normally the low pressure or return channel. The normallyhigh pressure and return channels 45 and 45' may alternate with eachother pressure-wise, depending upon the mode of operation of thetransmission, however, the shuttle spool 75 is always forced against thevalve seat on the high pressure side thus closing off the high pressurefrom the center of ducting core 13; on the oher hand the low pressureside is always in hydraulic communication with the center of the ductingcore 13 so that except for the pressure drop due to oil flow through thecommunication channeling, the low pressure or return prime channel 45 or45 will he at the same fluid pressure as the center of ducting core 13.By this device a changing pressure may be applied to the return fluidchannel 45 or 45 through channel 76, located at the center of inputshaft 1 (FIGURES l and 2), togdiscourage any tendency toward cavitationby pumping action of the driving and driven pumps and also to make upany leakage of oil from the inner housing of the transmission. Annulargroove 77 is provided in end piece 3 of the outer housing as a means ofeffecting hydraulic communication between an oil pressure source locatedexternally to the inner housing and channel 76 under conditions when theinner housing is rotating as well as when said inner housing is heldmotionless in respect to the outer housing as shown in FIGURES 1 and 2.

Referring to FIGURE 2 it will be apparent that spring loaded piston 78may be a pressure regulating device which will maintain a certainchanging pressure depending upon the spring characteristics and certainother factors when oil is forcibly circulated through channel 79 at thecenter of shaft 59. Oil flow would be through oil channel 76 fromannular groove '77 into the center of ducting core 13 from where itwould be transmitted through channel 79 located at the center of shaft59 to the face of piston 78, which would cause the piston to move towardthe right in FIGURE 2 until orifices 80 were exposed to the extentnecessary to pass the volume of oil being circulated. Oil would flowthrough orifices 89 into the hollow space in output shaft 2 partiallyoccupied by bearing assembly 60 from where said oil would then passthrough slots '64 and eventually along the lower side of the outerhousing where it would he returned to the sump through channel 81 in endpiece 3 (FIGURE 1). Circulation of oil in this manner may also provide ameans of transferring heat from the transmission proper to the sump orto a heat exchanger; however, under most conditions the power loss inthe transmission will he sufficiently low as to render the use of alarge capacity heat exchanger unnecessary.

The inner housing comprising parts aforementioned is journaled at theinput shaft end on hearings 82 and at the output shaft end on hearings83, said inner housing being disposed to rotate under certain conditionsas described herewith. When operating at an input-output forward speedratio of one to one, the reaction forces imparted to said inner housingthrough divider disk 15-15 are equal and opposite to the reaction forcesimparted to said inner housing through divider disk 18-18', said forceseffectively cancelling each other, thus permitting internal frictionalforces to rotate said inner housing without the necessity for fluidlocking. Assuming some leakage of fluid which bypasses the normalcirculatory paths provided in the transmission, output shaft 2 willexperience a slippage in speed relative to the speed of input shaft 1,in which case it is obvious that the inner housing cannot rotate atsynchronous speed with both input shaft 1 and output shaft 2. Dynamicforces acting on divider disk 15--15' of the driving pump and dividerdisk 1818 of the driven pump when the two are in asynchronous motioncause the inner housing to rotate at a speed intermediate between thespeeds of the input and output shafts due to the uniformity of pressurein directly communicative spaces.

Circulation of oil with the inner housing in synchronous rotation willbe near zero; however, this will not impair torque transfer inasmuch asthe torque transfer is a function of the net resultant hydraulic forcesacting on vanes 22 and 22A of the rotors 14 and 17 and not a function ofthe oil transfer.

When the input-output forward speed ratio is greater than one-to-one,and when the power source is supplying torque to input shaft 1, greatertorque is applied to driven pump rotor 17 then is applied to drivingpump rotor -14, resulting in a net reverse reaction force acting on theinner housing'through the respective divider disks 18-13 and -1616 whichtends to cause said inner housing to rotate in the opposite direction tothe rotation of the input and output shafts. The function ofover-running clutch 19, shown diagrammatically in FIG- URES 1 and 2, isto prevent the reverse rotation of the inner housing but to permit freeforward rotation.

-When the input-output reverse speed ratio is greater than one-to-one,and when the power source is supplying torque to input shaft 1, greatertorque'is applied to driven pump rotor 17 than is applied to drivingpump rotor 14, and the net reaction force acting on the inner housing inthis case tends to cause said inner housing to rotate in the forwarddirection which is undesirable inasmuch as said forward rotation wouldcause an effective reduction in the output-input torque ratio. Band 21,when caused to contract around the drum 20 with adequate force byhydraulic servo means or other suitable means, prevents forward rotationof the inner housing.

Under conditions when the impelling force is transmitted via outputshaft 2 to rotor 17, thence to rotor 14 and then via input shaft 1 tothe power source whichin this case becomes the load, as for example whena vehicle is descending a grade, and it is desired to increase the loadby forcing the power source to a higher speed by increasing thetransmission input-output speed ratio,

hydraulic forces tend to rotate the inner housing in,

the forward direction, which rotation if permitted to occur would lowerthe input-output speed ratio; therefore it is again desirable for theinner housing to be held against forward rotation by constriction of thehand 21 around the drum 20.

Shield 84 is provided for the purpose of reducing windage losses whenthe inner housing is rotating and is designed to cover thecircumferentially asymmetrical portion of the inner housing.

Possibilities for many different configurations of control systems existfor adapting a transmission of this type to automotive, industrial orother applications. FIG- URE 3-1 is a simplified schematic drawing of acontrol system which might be used to automatically control theinput-output speed ratio of the transmission in automotive vehicles as afunction of output shaft speed and torque demand, or input shaft speedunder certain conditions. The control system herein described is jointlyelectrical, mechanical and hydraulic in operation.

The hydraulic portion includes an oil pressure source and a follow upvalve assembly 101 containing a valve spool 102 which is axiallyslidable in valve sleeve MP3, said valve sleeve 103 being slidablymounted in cylinder 164. The lands 105 and 106 on the valve spool 102are of such width and spacing so as to completely occlude oil inletchannel 107 and oil return channel 108' when aligned 5.? shown in FIGURE31, but to permit circulation of oil from the pressure source 106through the control passages and back to the sump when said valve spool102 is displaced axially in either direction in respect to valve sleeve103. When valve spool 102 is moved to the left in FIGURE 31, oil underpressure from inlet channel 187 is permitted to enter control chamber Bthrough oil channel 109 and oil is permitted to return from controlchamber A to return channel 108 through oil passage 110 as indicated bythe solid arrows. By this means disk support assembly 1616' moves in thedirection of increasing forward pitch or decreasing reverse pitchcausing arm 68 of the ratio indicating linkage to move in the directionof the solid arrow. This action continues until inlet channel 107 andreturn chan nel 108 are realigned with lands 105 and 106 at which timeoil circulation into control chamber B and out of control chamber A isblocked. If valve spool 102 is moved to the right in FIGURE 31 oil ispermitted to enter control chamber A under pressure and to exit fromcontrol chamber B as indicated by the dotted arrows. This causes disksupport assembly 16-16 to move in the direction of decreasing forwardpitch or increasing reverse pitch causing arm 68 in this case to move inthe direction of the dotted arrow. This action continues until inletchannel 107 and return channel 108 are again realigned with lands 105and 106 at which time oil circulation into control chamber A and out ofcontrol chamber B is blocked. It is therefore apparent that when valvespool 102 is moved axially in either direction, that disk supportassembly 16-16 is caused by hydraulic means to follow said axialmovement of valve spool 102 by a corresponding angular amount in thedirection which will tend to maintain alignment of lands 105 and 106with inlet channel 107 and return channel 108.

An electric motor 111 is provided to effect axial movement of valvespool 102 through cooperation of the threaded parts of shaft 112 andspool shaft 113. Spool shaft 113 should be designed for axial movementonly. Motor 111 is controlled electrically by commutator-brush assembly114 containing four pairs of commutator-brush rotors as shownschematically in FIGURE 32; viewing from left to right, one pair forDrive, one pair for Neutral, one pair for Low and one pair for Reverse.Essentially the purpose of commutator-brush assembly 114 is to comparethe actual input-output speed ratio of the transmission at any instantwith what the ratio should be as a function of transmission output shaftspeed, torque demand, and selected speed range and to control motor 111in such a way as to cause said motor 111 to drive valve spool 102 to theposition which will correct any discrepancy.

Shaft 115 is geared at one end to spool shaft 113 and is directlycoupled to the Drive, Neutral, and Low brush arms or rotors and iscoupled through reverse gearing (not shown) to the Reverse brush arm andthereby causes angular movement of said brush arms when spool shaft 113moves axially. Shaft 116 is geared at one end to control rod 117 and iscoupled by suitable means to the Drive, Low, and Reverse commutationrotors so as to provide angular rotation in direct proportion to axialmovement of said control rod 117 up to certain predetermined limits foreach commutation rotor. For example, the limit of travel for the Drivecommutation rotor might be adjusted to correspond to an input-outputspeed ratio of one-to-one for the transmission, whereas the limit oftravel for the Low and Reverse commutation rotors might arbitrarily beadjusted to correspond to an input-output speed ratio of two-to-one forthe transmission. The Neutral commutation rotor is fixed in positioncorresponding to an input-output speed ratio of infinity, or zero outputfor the transmission driving pump.

Selector switch 118 may be designed to permit manual selection of anyone of the four switch contacts of the corresponding pairs ofcommutator-brush rotors and to provide electrical contact with the brusharm of the one selected. Only the selected pair of commutator-brushrotors may effect control of motor 111. In operation, when controlaction causes the ring F of the selected pair of commutator-brush rotorsto come in contact with the brush, a closed circuit is establishedbetween the positive and negative poles of the vehicle electrical systemvia the selector switch 118, brush arm, ring F and the winding of relayF. Relay F then establishes electrical connection to motor 111 to causesaid motor 111 to rotate in the direction which will cause valve spool102 to move to the left viewing FIGURE 31, until the gear 119 and shaft115 are rotated counter-clockwise by an amount which will cause thebrush arm to move the brush to the gap position (FIGURE 32) between therings F and R thus opening the electrical circuit which action causesthe motor to stop. When control action causes the ring R of the selectedpair of commutator-brush rotors to come in contact with the brush, aclosed circuit is established between the positive and negative poles ofthe vehicle electrical system via the selector switch 118, brush arm,ring R and the winding of relay R. Relay R then establishes electricalconnection to motor 111 to cause said motor 111 to rotate in thedirection which will cause valve spool 102 to move to the right viewingFIGURE 31 until the gear 119 and shaft 115 are rotated clockwise by anamount which will cause the brush arm to move the brush to the gapposition 120 between the rings F and R thus opening the electricalcircuit which action causes the motor to stop.

Since the Drive, Neutral, and Low brush arms are directly coupled toshaft 115, rotation of said brush arms is naturally in the samedirection as for shaft 115; however, inasmuch as the Reverse brush armis coupled to shaft 115 through reverse gearing, rotation of saidReverse brush arm is in opposition to the rotation of shaft 115. Thereverse gearing of the Reverse brush arm together with the oppositelyoriented F and R rings permits the Reverse commutation rotor to movecounterclockwise in respect to the Neutral reference position the sameas for the Drive and Low commutation rotors and to provide the sametransmission input-output ratio control characteristics in the Reversespeed as in the forward Low speed. The Neutral reference position forthe commutation rotors is defined by vertical orientation of the gapposition 120 as shown by the commutation rotor for Neutral in FIGURE 32.FIGURE 32 shows the selector switch 118 in the Low position in which aforward input-output speed ratio for the transmission is indicated by acounter-clockwise displacement of the Drive, Neutral, and Low brush armsin respect to the vertical Neutral reference. With the commutationrotors in the same position, it is apparent that if selector switch 118was put in the Reverse position that the servo action previouslydescribed would cause valve spool 102 to move to the positionrepresenting a reverse transmission speed ratio equal to the forwardspeed ratio as indicated in FIGURE 32 and the Reverse brush arm wouldthen be at the same counter-clockwise displacement angle in respect tothe Neutral reference as for the Drive, Neutral and Low brush arms shownin FIGURE 32 and that the latter brush arms would then be at the samerelative position as shown for the Reverse brush arm in FIGURE 32.

From the foregoing description it will be understood that input-outputspeed ratio control of the transmission in either forward or reverseoutput speeds is effected by control rod 117. As a means of causing saidcontrol rod 117 to move as a function of output shaft speed and torquedemand or input shaft speed, an output speed governor 121, a vacuummodulator 122 and an input speed governor 123 are provided.

Governor 121 may be a centrifugally operated device geared to orotherwise coupled to the transmission output shaft 2 and coupled tocontrol rod 117 as shown in FIGURE 31 to cause said control rod 117 tomove axially to the left, viewing FIGURE 31, as a function of outputshaft speed. Vacuum modulator 122 may be coupled to governor 121 in asuitable manner to oppose the action of said governor 121 as a functionof engine manifold vacuum and thus provide higher input-outputtransmission ratios with decreasing manifold vacuum or increasingly opencar-bureator throttle positions. With output shaft 2 motionless such aswould occur when a vehicle is at a standstill, governor shaft 124 may beset at such a position as to cause the transmission pump to be at thezero output or neutral position when spring 125 is holding control rod117 toward the right, in FIGURE 31, to the limit imposed by shoulder 126resting against the lip of sleeve 12'7.

Governor 123 may be a centrifugally operated device geared or otherwisecoupled to the transmission input shaft 1 and coupled to control rod 117as shown in FIG- URE 31 to cause said control rod 117 to move axially tothe left viewing FIGURE 31 as a function of input shaft s'pee Governor123 and associated coupling linkage may be designed such that at anominal input shaft speed, as for example the idling speed of a gasolineengine, shaft 128 will remain at the extreme right hand position, inFlGURE 31, but that at increasingly higher speeds of input shaft 1,shaft 128- will move to increasingly more left hand positions, viewingFIGURE 31, causing control rod 117 to move leftwardly with said shaft128 against the tension of spring 125 until a predetermined maximum lefthand position is reached, for instance corresponding to a transmissioninput-output ratio of four-to one, at which point shaft 128 will be heldagainst further leftwardly movement; however, control rod 117 will notbe restrained to further leftwardly movement resulting from leftwardlymovement of sleeve 127. This arrangement will permit smooth starts forvehicles at a standstill, when selector switch 118 is in the Drive, Low,or Reverse positions by permitting governor 122 to exercise control upto a predetermined limit, for instance corresponding to a vehicle roadspeed of five miles per hour, and permitting governor 121 in conjunctionwith a torque-demand sensing element such as vacuum modulator 122 toeffect transmission control at higher road speeds. As describedheretofore, bypass control piston 70 will prevent positive circulationof oil between the driving pump and driven pump of the transmission whenthe input-output speed ratio is above a predetermined value in forwardor reverse output speeds such as would be the case when an automotvevehicle is being accelerated from a standstill. I

A control means (not shown) of any conventional form may be provided forconstricting band brake 20 around drum 21 when either the Low or Reversepositions are selected on the selector switch .118. An interlock switchmay be provided to prevent starting of the engine except when disksupport assembly 1616 is at the zero pitch or zero output position.

The control system herein described is suited for automatic operation ofthe transmission for automotive application without any additional fluidcoupling interposed in the power train. The design of the transmissionand control system for use with a slippage type of fluid couplinginterposed between input shaft 1 and the engine crankshaft could dilferin several details.

An obvious constructional variation in the illustrated embodiment wouldbe the physical separation of the driving pump and driven pump with eachhaving its own housing and hydraulically connected by lengths of tubingor conduit. Such a variation would be equivalent to cutting thetransmission at line X-X ofFIGURE 1 into two parts, each part then beingadapted to make appropriate connections with the interconnecting highpressure and return tubes or conduits and removing the outer housing.For instance, by locating the driving pump near the engine and thedriven pump near the rear axle of an automobile, the need for theconventional drive shaft and associated universal joints would beeliminated thus making possible a fiat floor design for automobiles. Byproviding two driven pumps or one for each rear wheel, the conventionaldilferetnial gearingcould be also eliminated.

A further possibility would be to integrally combine the driven pumpswith the wheels.

C-onstructional variations such as four wheel drive and front wheeldrivein lieu of rear wheel drive could more readily be implemented with ahydraulic system than with conventional mechanical power coupling means.The possibility also exists for combining the braking means with thepropelling means.

While there is given above a certain specific example of this inventionand its application in practical use, it should be understood that thisis not intended to be exhaustive or to be limiting of the invention. Onthe contrary, this illustration and explanation herein are given inorder to acquaint others skilled in :the art with this invention andthe. principles thereof and a suitable manner of its application inpractical use, so that others skilled in the art may be enabled tomodify the invention and to adapt and apply it in numerous forms each asmay be best suited to the requirement of a particular use.

I claim:

1. A variable speed transmission of the hydraulic type comprising ahousing having therein a driving pump chamber and a driven pump chamber;input and output shafts respectively mounted in said chambers; drivingand driven pump assemblies respectively mounted in said chambers, saidassemblies including a rotor connected to said input and output shaftsand a plurality of radial vanes mounted on said rotors; a divider disksupporting assembly mounted in said housing and having therein anannular groove, a :divider disk having a plurality of radial slotstherein mounted in said groove and adapted to intermesh with said vanes,a plurality of peripheral chambers on each side of said divider disk inthe portion thereof disposed in said annular groove, each of saidperipheral chambers being connected by a duct to a corresponding pumpchamber formed between adjacent vanes on the opposite sides of saiddivider disk, said disk support assembly being pivotally mounted about adiameter thereof in said housing about said driving pump assembly; aducting core extending from said driving pump chamber to said drivenpump chamber, said ducting core having at least a pair of channels outtherein to sequentially connect together corresponding pump chambers ofsaid rotor assembly whereby oil may flow from one to the other; andcontrol means for varying the pitch of said driving pump divider diskassembly.

2. A device as described in claim 1 wherein the outer edge of each vaneof said driving pump rotor has a slot cut therein and said divider diskhas at the center of each of said radial slots a cooperating pin adaptedto engage in said slot, said slot being contoured to maintain said diskin proper phase relationship to said rotor.

3. A variable speed transmission of'the hydraulic type comprising ahousing having therein a driving pump chamber and a driven pump chamber;input and output shafts respectively mounted in said chambers; drivingand driven pump assemblies respectively mounted in said chambers, saidassemblies including a rotor connected to said input and output shafts,and a plurality of v-anes mounted on said rotor; a divider disksupporting assembly mounted in said housing and having therein anannular .groove; a divider disk mounted in said groove; a plurality ofradial slots cut in said disk and adapted to intermesh with said vanesmounted on said rotor; said disk supporting assembly forming with saidrotor and disk a plurality of pump chambers on each side of said disk; aplurality of peripheral chambers on each side of said divider disk inthe portion thereof disposed in said annular groove, each of saidperipheral chambers being connected by a duct to a corresponding pumpchamber; said disk support assembly in said driving pump assembly beingpivotal-1y mounted in said housing about a diameter thereof disposed atright angles to the axis of said housing, a ducting core extending fromsaid driving pump chamber to said driven pump chamber; a plurality ofchannels cut in said ducting core to sequentially connect togethercorresponding pump chambers of said rotor assemblies whereby oil mayflow from one to the other; control means for varying the pitch of saiddriving pump divider-disk assembly; an outer casing surrounding saidhousing, driving and driven pump assemblies; bearing means within saidouter casing to permit rotation of said driving and driven pumpassemblies therein; overrunning clutch means mounted between said outercasing and said housing to prevent reverse rotation thereof and handbrake means connected between said outer casing and housing toselectively prevent forward rotation of said housing.

4. A variable speed transmission of the hydraulic type comprising ahousing having therein a driving pump chamber and a driven pump chamber;input and output shafts respectively mounted in said chambers; drivingand driven pump assemblies respectively mounted in said chambers, saidassemblies including a rotor having six radial vanes, connected to saidinput and output shafts, and a divider disk having six radial slotstherein adapted to intermesh with said vanes mounted on said rotor; adivider disk supporting assembly having therein an annular grooveadapted to receive the periphery of said divider disk, said disksupporting assembly forming about said rotor and disk an inner chamberWithin said housing, having twelve pump chambers; six peripheralchambers on each side of said divider disk in the portion thereofdisposed in said support-ing assembly annular groove, each of saidperipheral chambers being connected by a duct to a pump chamber on theopposite side of said divider disk; said driving pump assembly alsohaving said disk support assembly pivotally mounted in said housingabout a diameter thereof, a ducting core positioned within said drivingrotor assembly and extending from said driving pump chamber to saiddriven pump chamber, said ducting core having at least a pair ofchannels out therein to sequentially connect together corresponding pumpchambers of said rotor assemblies whereby oil may flow from one to theother; control means for varying the pitch of said driving pump dividerdisk assembly; an outer casing surrounding said housing, driving anddriven pump assemblies; bearing means fixed in said outer casing andhaving journaled therein said driving and driven pump assemblies;overrunning clutch means operatively mounted between said outer casingand said housing to prevent rotation thereof in one direction and handbrake means operatively connected between said casing and housing toselectively prevent rotation of said housing in the other direction.

5. A variable speed transmission of the hydraulic type comprising ahousing having therein a driving pump chamber and a driven pump chamber;input and output shafts mounted respectively in said chambers; drivingand driven pump assemblies respectively mounted in said chambers, saidassemblies comprising a rotor having six radial vanes mounted about andconnected to said input and output shafts, a divider disk having sixradial slots therein adapted to intermesh with said vanes mounted onsaid rotor, a divider disk supporting assembly mounted in said housingand having therein an annular groove adapted to receive the periphery ofsaid divider disk, said disk supporting assembly forming about saidrotor and disk an inner chamber Within said housing, s-ix peripheralchambers formed on each side of said divider disk in the portion thereofdisposed in said supporting assembly annular groove, each of saidperipheral chambers being connected by a duct and valve mechanism to acorresponding pump chamber formed between adjacent vanes on the oppositeside of said divider disk; said disk support assembly in said drivingpump assembly being pivotally mounted in said housing about a diameterthereof disposed at right angles to the axis of said housing beingcontoured on the outer surface to form an oil sealing contact with theinner surface of said housing; a pair of vanes mounted in the plane ofthe axis of said support assembly and said shaft to form oil sealingchambers above and below the axis of said support disk assembly on atleast one side thereof, duct means for introducing and withdrawing oilfrom said chambers formed between said housing and said supporting diskassembly whereby the pitch of said assembly may be controlled, a ductingcore positioned within the shafts of said rotor assemblies and extendingfrom said driving pump chamber to said driven pump chamber, said ductingcore having at least a pair of channels cut therein to sequentiallyconnect together corresponding pump chambers of said rotor assemblieswhereby oil may fiow from one to the other; control means for varyingthe pitch of said driving pump divider disk assembly; angle indicatingmeans geared to said driving pump disk supporting assembly and extendingthrough said ducting core member to a movable collar on said outputshaft, lever means interconnecting said collar to said control means; anouter casing surrounding said housing, driving and driven pumpassemblies, bearing means mounted in said outer casing having saiddriving and driven pump assemblies journalled therein; overrunningclutch means mounted between said outer casing and said housing toprevent reverse rotation thereof and band brake means connected betweensaid housing and inner chamber to selectively prevent forward rotationof said housing.

6. The device of claim 5 wherein said angle indicating means includes asleeve, an annular channel in said sleeve, a plurality of orifices inthe high and low pressure channels of said ducting core, said sleevebeing positioned about said orifices when said driving pump dividerdiskassembly is near the zero pitch position to at least partiallyinterconnect through said channel said orifices whereby a slightbypassing of fluid is obtained to prevent stalling of the prime powersource when starting the driven pump under heavy loads.

7. A device as described in claim 5 wherein said ducting core hastherein a valve and port assembly interconnecting opposite pressurechannels thereof to the center of said ducting core, said valvemechanism being arranged to close oif the port leading to the highpressure side of said ducting core at any given time whereby oil may beadded to the low pressure side of said ducting core through the centerthereof.

8. A variable speed transmission of the hydraulic type comprising ahousing having therein a driving pump chamber and a driven pump chamber;input and output shafts respectively mounted in said chambers; drivingand driven pump assemblies respectively mounted in said chambers, saidassemblies including a rotor connected to said input and output shaftsand a pluarlity of radial vanes mounted on said rotors; a divider disksupporting assembly mounted in said housing and having therein anannular groove, a divider disk having a plurality of radial slotstherein mounted in said groove and adapted to intermesh with said vanes,said disk support assembly being pivotally mounted about a diameterthereof in said housing about said driving pump assembly; a ducting coreextending from said driving pump chamber to said idriven pump chamber,said ducting core having at least a pair of channels cut therein tosequentially connect together corresponding pump chambers of said rotorassembly whereby oil may flow from one to the other; and control meansfor varying the pitch of said driving pump divider disk assembly.

9. In a hydraulic variable displacement transmission of the type havinga driving pump connected to a prime power source and a driven pumpconnected to the source to be moved, control means comprising a pitchcontrol mechanism connected to said driving pump means, said pitchcontrol mechanism having forward, neutral, and reverse positions; a feedback loop for indicating the pitch of said driving pump in said pitchcontrol mechanism; motor means operatively connected to said pitchcontrol for moving said pitch control to the desired position; acommutator brush assembly having a plurality of commutation stators androtors; a load demand sensing element connected to said commutator brushassembly; switch means for selectively connecting to the desiredcommutation rotor; battery means connected through said switch means tosaid commutation rotors and load demand sensing elements connected tosaid commutator brush assembly whereby variations in load demand orswitch control will cause said pitch control mechanism to vary the pitchof said driving pump so that the desired output will be delivered to theoutput shaft.

References Cited in the file of this patent UNITED STATES PATENTS951,064 Erickson Mar. 1, 1910 Erickson Mar. 12, Cuny May 13, Haines May4, McGill July 13, Dodge Mar. 13, Jakobsen Nov. 18, Kraft June 8, CunyOct. 12, Paulsrn-eier Oct. 12, Marshall Apr. 1,

FOREIGN PATENTS France June 4, Germany May 29,

1. A VARIABLE SPEED TRANSMISSION OF THE HYDRAULIC TYPE COMPRISING AHOUSING HAVING THEREIN A DRIVING PUMP CHAMBER AND A DRIVEN PUMP CHAMBER;INPUT AND OUTPUT SHAFTS RESPECTIVELY MOUNTED IN SAID CHAMBERS; DRIVINGAND DRIVEN PUMP ASSEMBLIES RESPECTIVELY MOUNTED IN SAID CHAMBERS, SAIDASSEMBLIES INCLUDING A ROTOR CONNECTED TO SAID INPUT AND OUTPUT SHAFTSAND A PLURALITY OF RADIAL VANES MOUNTED ON SAID ROTORS; A DIVIDER DISKSUPPORTING ASSEMBLY MOUNTED IN SAID HOUSING AND HAVING THEREIN ANANNULAR GROOVE, A DIVIDER DISK HAVING A PLURALITY OF RADIAL SLOTSTHEREIN MOUNTED IN SAID GROOVE AND ADAPTED TO INTERMESH WITH SAID VANES,A PLURALITY OF PERIPHERAL CHAMBERS ON EACH SIDE OF SAID DIVIDER DISK INTHE PORTION THEREOF DISPOSED IN SAID ANNULAR GROOVE, EACH OF SAIDPERIPHERAL CHAMBERS BEING CONNECTED BY A DUCT TO A CORRESPONDING